This invention relates to vehicle suspension systems, and more particularly to a method and apparatus for a low power active suspension system for dynamic leveling of a motor vehicle.
In recent years, substantial interest has grown in motor vehicle suspension systems which can offer improved comfort and road holding over the performance offered by conventional passive suspension systems. In general, such improvements are achieved by utilization of an "intelligent" suspension system capable of electronically controlling the suspension forces generated by hydraulic actuators provided at each corner of the motor vehicle.
Suspension systems are provided to filter or "isolate" the vehicle body from vertical road surface irregularities as well as to control body and wheel motion. In addition, it is desirable that the suspension system maintain an average vehicle attitude to promote improved platform stability during maneuvering. The classic passive suspension system includes a spring and a damping device in parallel which are located between the sprung mass (vehicle body) and the unsprung mass (wheel and axles).
Hydraulic actuators, such as shock absorbers and/or struts are used in conjunction with conventional passive suspension systems to absorb unwanted vibration which occurs during driving. To absorb this unwanted vibration, the hydraulic actuators are automobile. A piston is located within the actuator and is connected to the body of the automobile through a piston rod. Because the piston is able to limit the flow of damping fluid within the working chamber of the actuator when the actuator is telescopically displaced, the actuator is able to produce a damping force which counteracts the vibration which would otherwise be directly transmitted from the suspension to the vehicle body. The greater the degree to which the flow of damping fluid within the working chamber is restricted by the piston, the greater the damping forces which are generated by the actuator.
In selecting the amount of damping that an actuator is to provide, three vehicle performance characteristics are often considered: ride comfort, vehicle handling and road holding ability. Ride comfort is often a function of the spring constant of the main suspension springs of the vehicle, as well as the spring constant of the seat, tires, and the actuator. Vehicle handling is related to the variation in the vehicle's attitude (i.e., roll, pitch and yaw). For optimum vehicle handling, relatively large damping forces are required to avoid excessively rapid variation in the vehicle's attitude during cornering, acceleration, and deceleration. Road holding ability is generally a function of the amount of contact between the tires and the ground. To optimize road holding ability, large damping forces are required when driving on irregular surfaces to prevent loss of contact between the wheels and the ground for an excessive period of time.
To optimize ride comfort, vehicle handling, and road holding ability, it is generally desirable to have the damping forces generated by the actuator be responsive to the input frequency from the road. When the input frequency from the road is approximately equal to the natural frequency of the body of the automobile (e.g., approximately between 0-2 Hz), it is generally desirable to have the actuator provide large damping forces to avoid excessively rapid variation of the vehicle's attitude during cornering, acceleration and deceleration. When the input frequency from the road is between 2-10 Hz, it is generally desirable to have the actuator provide low damping forces so as to produce a smooth ride and allow the wheels to follow changes in road elevation. When the input frequency from the road is approximately equal to the natural frequency of the automobile suspension (i.e., approximately 10-15 Hz), it may be desirable to have relatively low damping forces to provide a smooth ride, while providing sufficiently high damping forces so as to prevent excessive loss of contact between the wheels and the ground.
To obtain the desired damping characteristics over a wide range of driving conditions, efforts have been made toward the development of continuously variable damper valving. Such variable damper valving is able to adjust or vary the damping characteristics of a hydraulic actuator in response to rapid variations which occur during cornering, acceleration, and deceleration.
Recent advancements in sensor and microprocessor technology have created substantial interest in the application of "active" suspension systems in motor vehicles. An "active" suspension is defined as a suspension system with real time control of body and wheel motion. An actuator or force generator replaces or is added to the conventional passive suspension member. The dynamic behavior of a motor vehicle can be dramatically modified through application of an "active" suspension system.
Referring to FIG. 1, a diagrammatical illustration of a simplified control scheme for an active suspension is shown to include a closed loop control system 10 wherein the behavior of vehicle 12 is measured by multiple sensors 14 (i.e. accelerometers, gyroscope, potentiometers). A microprocessor-based controller 16 processes the sensed signals and calculates a demanded suspension force "U" for each actuator 18. Actuators 18 are provided at each corner of the vehicle and are independently controlled. Vehicular Inputs 20, which effect the dynamic behavior of vehicle 12, include disturbances such as road undulations, inertia forces generated during cornering and breaking, and aerodynamic forces. The active suspension is designed to compensate for these disturbances while providing improved ride isolation, improved road handling, reduced pitch and roll angles, control of lateral load transfer between front and rear axles, and control of the vehicle ride height above the road.
In general, active suspension systems are technically complex due to the demand for fast response and accuracy from the hydraulic system, the actuators and the control system. Heretobefore, application of active suspension systems in motor vehicles has been limited due to the excessive cost associated with the complex hardware and controls.
The demanded force signal "U" can be described as a function of state variables describing the vertical motion of vehicle 12 and of the road conditions. FIG. 2 illustrates a simplified 1/4 model of vehicle 12 having two degrees of freedom (DOF) which is described by four state space variables. The state variables include body displacement (X.sub.b), body velocity (X.sub.b), wheel displacement (X.sub.w) and velocity (X.sub.w). Inputs from road 22 are identified as (X.sub.road). An active suspension actuator 18 is diagrammatically illustrated as suspended between body 24 and wheel 26 relative to road 22 and in parallel with a passive suspension member 28. Control law for the calculation of the demanded force "U" for the single active actuator 26 shown is: EQU U=-G.multidot.X (1) EQU where: EQU U=g.sub.1 (X.sub.b)+g.sub.2 (X.sub.b)+g.sub.3 (X.sub.w) +g.sub.4 (X.sub.w)+g.sub.5 (X.sub.road) (2)
Gains g.sub.1 and g.sub.3 represent inertial damping and g.sub.2 and g.sub.4 reflect inertial stiffness for the body and wheel, respectively. Feedback of the road g.sub.5 is provided to tie wheel 26 to the road 22. Various techniques are known for defining a feedback gain matrix (i.e., Linear Quadratic Gaussian) to best fulfill specific design objectives associated with dynamically controlling the behavior of vehicle 12.
In general, active suspension systems can be divided into three basic classes; "semi-active", "low power" active, and "fully" active. A primary difference between each of the aforementioned active suspension classes relates to the hardware requirements to be hereinafter described. In general, "low power" active and "fully" active suspension systems each use a central hydraulic system to power each suspension actuator. "Semi-active" systems, however, only have the ability to dissipate power by varying the damped resistance to motion.
Semi-active damping systems are based on continuous, real time modulation of damping rates, separately in compression and extension. Semi-active control strategy is based on the principle that a suspension actuator is only capable of dissipating power. Referring to FIGS. 3 and 4, an actuator 32 for a semi-active suspension system is illustrated. In practical terms, semi-active actuator 32 is equipped with infinitely variable camper valving 34. A self-contained source of damping fluid is confined within working chambers 36 and 38 provided on opposite sides of piston 40. Compared to a conventional passive damper, semi-active actuator 32 can dissipate power to provide a demanded force "U". However, the control strategy for semi-active suspension systems must be simplified from that of equation (2) to include only velocity of the vehicle body (X.sub.b) and the relative suspension velocity (X.sub.w -X.sub.b).
For power dissipation: EQU U=g.sub.1 (X.sub.b)+g.sub.2 (X.sub.w -X.sub.b) (Quandrants I and III)(3)
For power demand: EQU U=O (Quadrants II and IV) (4)
In "ideal" semi-active systems, the demanded force "U" is set to zero (U=O) if the relative velocity (X.sub.w -X.sub.b) is of the opposite sign and/or the demanded force "U" is out of "reach" for actuator 32. These conditions are reflected as Quadrants II and IV of the Force "U" versus relative velocity (X.sub.w -X.sub.b) graph of FIG. 4. Therefore, as is readily apparent, semi-active systems are capable of dissipating power (Quadrant I and III) but are not capable of generating all demanded force requirements (i.e. Quadrant II and IV) associated with dynamically controlling vehicle behavior.
Modernly, "low power" active systems have heretobefore been used almost exclusively on race cars. Unfortunately, the frequency response of low power active systems is restricted to a undesirably narrow frequency range (0-5 H.sub.z) as well as a significantly lower power output than required for "fully " active systems. FIG. 5 schematically illustrates a single-acting hydro-pneumatic actuator 42 applicable for low power active suspension systems. Because of the low frequency response of actuator 42, the control strategy of Equation (2) is limited to a narrower frequency range which necessitates exclusion of high frequency control of the wheel. Therefore the control equation is defined as: EQU U=g.sub.1 (X.sub.b)+g.sub.2 (X.sub.b) (5)
Low power systems have a central pump 50 supplying high pressure to each corner actuator 42. Hydro-pneumatic actuator 42 communicates with an accumulator 44 through a fixed flow damper valve 46 for damping the suspension motion. A servo valve 48 is either supplying actuator 42 with high pressure fluid provided from a central pump 50 or emptying actuator 42 of fluid which is then returned to a central reservoir 52. The suspension stiffness is supplied by a passive spring member (not shown) and the pressure in accumulator 44. In low power active systems, hydraulic damping fluid is selectively controlled to flow between central reservoir 52 and actuator 42 in response to changes in the vehicle's attitude.
"Fully" active suspension systems are designed to control both wheel and body motion over the entire frequency range (approximately 0-30 H.sub.z). A "fully" active suspension actuator 60 is schematically illustrated in FIG. 6. Specifically, the central hydraulic system (i.e., pump 50 and reservoir tank 52) is basically the same as the one used for the low power active but designed to provide higher flow rates for increased peak power consumption. Actuator 60 is basically a double-acting hydraulic device having an upper working chamber 62 and a lower working chamber 64 defined by opposite sides of piston 66. A four port servo valve 68 connects each working chamber 62 and 64 to the high supply pressure source (pump 50) or back to central reservoir 52. The controlling formula, that is, equation (2) can be used to its full merits to include control to both vehicle body and wheel motion.
However, the current costs associated with "fully" active suspension systems are extremely high, primarily because of the fast frequency response requirements of servo valve 68 and the high peak flow rates associated with the central hydraulic system. Actuator 60 is not equipped with damper valving through piston 66 whereby system power requirements are extremely high to generate movement of the piston and rod directly from the pressure difference generated across piston 66. Therefore, high system pressure requirements, large peak power and peak fluid flow requirements are associated with conventional "fully" active suspension systems. Furthermore, the duty cycle of such a "fully" active suspension system is almost continuous due to constant modulated actuation of servo valve 68 to control fluid flow to and from actuator 60. Accordingly, there is a need to reduce the complexity, size of the components, system cost, peak system working pressure, and peak power requirements, while effectively generating superior operational durability characteristics.
It is to be understood that the terms "height", "distance", "attitude", and derivatives thereof are used interchangeable herein as well as throughout the automotive art, as referring to the magnitude of spacing between a vehicle's sprung and unsprung portions (e.g., between a vehicle frame and its associated axles or independent wheels).